There are a wide variety of different types of bearings. Common types of bearings include devices that support the rotating shaft in rotating machinery, allowing for continuous rotation. Common examples of this type of bearing include rolling element bearings, oil lubricated bearings, thrust bearings, and active magnetic bearings. Other types of bearings only allow for a restricted range of motion of a supported component or structure. Two common examples of this type of bearing include squeeze film dampers used in rotating machinery and seismic isolator bearings used in large building foundation systems and bridge supports for protection from earthquake motion. Some bearing assemblies, such as compliant foil bearings, themselves contain a bearing element as an integral subcomponent.
The idea of using a relatively compliant bearing to support a relatively stiff bearing has a long history of improving overall system performance. For example, in 1926, Trumpler (U.S. Pat. No. 1,595,744) described a series of flexible support systems applied both to rolling element (ball) bearings, as well as hydrodynamic radial bearings. These flexible support systems were described as being soft enough to allow the shaft to rotate around its mass center, thus accommodating variable unbalance distributions. Trumpler's flexible mount is described as also using oil filled spaces which are squeezed as the shaft orbit to provide damping. Over time, support bearings using oil filled clearance spaces used for rotating machinery have come to be known as “squeeze-film dampers.”
This idea of a soft mount with squeeze film damping is now very widely used [Vance, 1988], most notably in nearly every modern aircraft gas turbine engine. The low stiffness of squeeze film dampers allow engineers to place rotor-bearing system natural frequencies (often referred to as “critical speeds”) at frequencies well below the normal running speed range, while simultaneously providing enough damping to ensure that dangerous vibration amplitudes do not occur as the shaft rotating speed traverses through these frequencies. Squeeze film dampers are also sometimes used to reduce unstable whirling motions in high speed rotating machinery.
Conventional oil lubricated squeeze film dampers have multiple drawbacks. One major issue is their limited operating temperature range. Conventional oil lubricants cannot be used at temperatures much over about 450 degrees F. without experiencing chemical degradation or potentially catching fire. At very low temperatures, oils can become extremely viscous and lose the ability to flow easily. Another drawback to conventional squeeze film dampers in some applications is the need for a constant supply of lubricating oil. Many smaller machines that would benefit from squeeze film dampers use grease lubricated bearings. Thus there is no oil supply to allow a conventional oil-lubricated squeeze film damper to be employed.
Salahi, Heshmat and Walton (2007) discuss using a frictionally damped metallic spring element as a compliant support bearing as an alternative to the conventional squeeze film damper. They report that a frictionally damped support structure offers high levels of performance in this application without the drawbacks of oil. However, the compliant support element they present contains numerous small, precision formed parts which must be precisely assembled with spot welds. This construction is costly and difficult to produce.
The frictionally damped bearing element that Salahi et al. propose as a squeeze film damper replacement was originally developed as a flexible support bearing integrated directly into a gas lubricated compliant hydrodynamic bearing assembly. Gas lubricated hydrodynamic bearings with an integrated compliant bearing element have been around since at least the mid 1960's. One early example is a 1965 patent issued to Marley (U.S. Pat. No. 3,215,480). This patent describes a gas lubricated hydrodynamic bearing which uses a concentric array of flexible thin foil strips supported at the ends, and loaded in the center as the flexible non-rotating bearing component. During operation, hydrodynamic pressure is generated in the lubricant film (typically air) between the shaft and top foil(s) as the shaft rotates at high speed in the direction of a converging clearance between the top foil(s) and the shaft. The hydrodynamic pressure acts to separate the shaft and flexible bearing surface, and support the weight of the rotating component.
Since their development in the 1960's, the complexity and design sophistication of foil bearing designs have steady increased in the pursuit of increased performance with regards to load capacity and the bearing's effects on rotor-bearing system dynamics.
One of the first advancements in the design of foil bearing integral flexible support bearing elements was to use multiple spring elements along the length of the bearing assembly. Dellacorte and Valco (2000), for example, shows two flexible support elements which have an axially split strip of corrugated bump foils. This axially split construction improves the bearing's ability to handle misalignment. It also makes it possible to vary the support element stiffness in the axial direction. For example, it is thought to be advantageous to have a lower stiffness near the top foil edges. Other foil bearing developers have proposed using variable pitch bumps and/or heights in the flexible element to achieve circumferentially varying stiffness. An example of this approach is described in U.S. Pat. No. 4,262,975. This patent also presents axially split bumps and an additional “stiffener” foil between the top foil and the bump foils. Although the bearing using this flexible support assembly is described as having very high levels of performance, the support assembly is quite complex. As with most bump foil based configurations, the disclosed configuration relies on interconnected arch forces to support the load, it thus has the drawback that there is interaction between the forces applied to one bump, and deflections of other bumps in the strip. There is no localized load-bearing capability. For long strips of bumps, such as disclosed in FIG. 3, friction forces can build up and prevent some of the bumps from sliding. This reduces the ability of the bump strip to dissipate vibrational energy through friction.
U.S. Pat. No. 4,300,806 describes a configuration using staged multiple nested stiffness elements with gaps to achieve a nonlinear stiffness characteristic that hardens with increasing load. This approach is described as improving overall bearing performance by having a low stiffness support element at low speeds and loads, then an increasing stiffness support element at higher speeds and loads. This flexible support element is complex to produce, since it has numerous precision formed subcomponents.
U.S. Pat. No. 4,315,359 describes a foil thrust bearing compliant support that uses a one dimensional circumferential array of “flaps” formed towards the top foil, which is cooperatively engaged with a disk-like feature on the rotating shaft. It is indicated that an object of the patent is to provide a compliant fluid film thrust bearing which permits the top foil to deform locally, without affecting the pad height at other locations. It also suggests that the required manufacturing tooling is economical to produce. Two significant drawbacks of the proposed bearing are that it requires welded components, and the “flaps” are not split into multiple sections or two-dimensional arrays to more readily accommodate misalignment and provide for localized load-bearing capabilities.
U.S. Pat. No. 4,699,523 describes a radial foil bearing wherein a bearing foil is retained in an axial profiled slot machined into the wall of the bearing housing. This approach avoids the need to use spot welds that are often used for this purpose.
U.S. Pat. No. 5,116,143 describes a radial bearing with compliant support element stiffness for a radial foil bearing which uses a one-dimensional circumferential array of “curvilinear support beams” that vary in width and pitch to optimize the support stiffness distribution. The support beams are envisioned as extending nearly the entire axial length of the bearing, precluding the ability to provide localized load-bearing capabilities. Small cutouts at the ends of the beams are used to provide a reduced stiffness at the edges of the bearings. As noted previously, a drawback of this construction is that a single spring element extending the entire length of the bearing has a limited ability to accommodate misalignment or local deformations in the bearing.
U.S. Pat. No. 5,427,455 presents a compliant support element with small cantilever beams which are created when a foil with specially shaped cutouts forced into a circular shape to fit into a bearing housing. This compliant support element and the resulting radial foil bearing partially address some of the drawbacks identified above. The advantages include:
1. It is possible to vary to the support stiffness both axially and circumferentially.
2. The support element can readily accommodate misalignment and local deformations, since multiple spring features are used axially.
3. No welds are required to assemble the bearing.
4. The spring element is economical to produce.
However, this compliant support element has some important drawbacks. Most notable is that it cannot be used for anything other than a radial bearing, because the cantilever beams are formed only when the foil layer is fixed into a cylindrical shape. A generally planar thrust bearing, for example, is not feasible. Even from the perspective of a radial bearing, it does not allow the designer to vary the angle and distance between the cantilever beam spring and the stationary counter-surface which supports the tabs in any meaningful way, because the cantilever beam springs are created as a side effect of the compliant element being bent into a circular shape, rather than as an explicitly formed profile. Thus, the configuration has a limited ability to provide localized load support and/or stiffness. The need to force the support element structure into a circular shape also tends to result in a bearing with either an undesirably large amount of radial force (preload) between the shaft and the top foil, and/or requiring retainers which prevent the spring element from springing back towards a flat shape to be included in the bearing housing's inner profile. It is also difficult to produce a reliable, practical bearing with less than three pads. This limits the achievable load capacity for this configuration, because three-pad compliant foil bearings tend to have less load capacity than one or two pad designs.
U.S. Pat. No. 5,938,341 presents a generally planar foil thrust bearing support which uses a one-dimensional circumferential array of formed spring tabs to apply axial preload to a bearing. The objective of the assembly is to be a preload spring, rather than a support spring for the operating surface of the bearing. The use of the preload springs as a means to dissipate vibrational energy is not discussed in this patent. The patent does not envision a large number of spring tabs, or multi-dimensional arrays of tabs, as would be required to support the top foil surface of a compliant foil bearing. The patent is also silent on varying tab shape, height, bend angle(s), etc.
U.S. Pat. No. 8,360,645 presents a foil bearing support structure which is quite similar to U.S. Pat. No. 5,116,143, except that the cantilever beams or flaps are formed by folding/pleating the foil, rather than with cutouts and tabs that are bent into the required shape. It suffers from the same disadvantages as U.S. Pat. No. 5,116,143.
DellaCorte and Valco (2000) summarized much of the progress in foil bearing compliant support design by dividing radial foil bearings into three groups:
1. “Generation I” bearings, which have a more or less uniform support stiffness distribution.
2. “Generation II” bearings, which have support stiffnesses that vary in either the circumferential or axial direction.
3. “Generation III” bearings, which have support stiffnesses that can vary in both the circumferential and axial directions, as well as possibly nonlinear stiffness with load characteristics.
Generation I bearings often have the lowest performance, while Generation III bearings often have the highest performance. The major difference between these three groupings is the flexible support element(s). In general, contemporary Generation III bearings are more complex with more subcomponents and would often be expected to be more expensive to design and produce.
Thus, although a large number of compliant supports for foil bearings have been proposed, contemporary compliant supports suffer from several disadvantages, including:
1. Many configurations use compliant elements which extend the full length of the bearing top foil. Full length compliant elements limit the compliant support element's ability to accommodate misalignment or localized deformations of the bearing or rotating shaft.
2. Most configurations have limitations with regards to independently and arbitrarily varying the support element stiffness spatial distribution.
3. Some configurations are only suitable for either generally circular or generally planar bearings, but not both.
4. As they are typically implemented and produced for radial foil bearings, many configurations have constant height bumps. This approach limits the bearing designer's ability to create an initial bearing clearance profile which has one or more converging wedge regions for a centered shaft. This preload is well known to often improve stability in rigid surface (fixed geometry) hydrodynamic radial bearings. The preload may also be beneficial in compliant surface bearings.
5. Most configurations require numerous parts that must be precisely formed, and accurately assembled with tolerances on the order of 0.0001 inches, often with multiple spot welds. The large number of parts, need for spot-welding, and required assembly tolerances tend to make these designs expensive and inherently unsuitable for high volume production.